Slidable vane rotary compressor



March 20, 1962 R. EMANUEL SLIDABLE VANE ROTARY COMPRESSOR 3 Sheets-Sheet 1 Filed March 13, 1961 March 20, 1962 R. EMANUEL 3,026,021

SLIDABLE VANE ROTARY COMPRESSOR Filed March 15, 1961 3 Sheets-Sheet 2 March 20, 1962 R. EMANUEL SLIDABLE VANE ROTARY COMPRESSOR 3 Sheets-Sheet 3 Filed March 15, 1961 United States Patent Ofiice Patented Mar. 20, 1962 3,026,021 SLIDABLE VANE ROTARY COMPRESSOR Roberto Emanuel, Turin, Italy, assignor to Soc. Acc- Emanuel di G. e R. Emanuel & C., Turin, Italy Filed Mar. 13, 1961, Ser. No. 95,076 Claims priority, application Italy Mar. 31, 1960 1 Claim. (Cl. 230-149 This invention relates to rotary compressors having slidable vanes, comprising in combination: a stator, a cylinder rotatable in the stator, a cylindrical rotor rotatable in the cylinder about an axis parallel with but offset from the cylinder axis, said rotor being constantly tangential to the cylinder along a contact generatrix, a number n of vanes capable of sliding in an at least ap proximately radial direction in the rotor and including a driving vane pivoted to the cylinder about an axis parallel with the axes of the cylinder and rotor, whereby the rotor and cylinder are coupled for rotation in synchronism in the stator, slots or similar fluid-transfer apertures in the cylinder between each pair of vanes, and induction and delivery ports in the stator with which said slots are successively connected on rotation of the rotor and cylinder.

A compressor as defined above will be briefly referred to hereafter as a compressor of the type specified.

The heaviest drawback of compressors of the type specifier known to date resides in a low efiiciency and varying delivery pressure. More particularly, the mechanical energy expended for driving the compressor largely exceeds the compression work, moreover the delivery pressure is not constant but undergoes periodical fluctuations departing by up to 50-60% up and down with respect of a nominal value.

This results in a number of further drawbacks, such as, for instance, low compression ratio, vibrations, noise, non-uniform wear of moving parts, etc.

It has now been found that all above mentioned drawbacks can be easily removed Without making use of any additional members or devices or introducing any appreciable complication in structure or in the manufacturing process of the compressor. In this connection the invention starts from the concept that angular distribution of the vanes and slots should not be determined according to a generic criterion of symmetry or the like, but should necessarily take care of a large number of factors some of which were not considered heretofore in the art.

According to this invention a slidable vane rotary compressor of the type specified is characterized in that, considering the driving vane on the tangency generatrix and taking as a reference the radial plane extending through the said generatrix and the axis of the cylinder, the azimuth angles oa of the remaining vanes around the rotor axis and the azimuth angles of their respective slots around the cylinder axis, measured in the direction of the rotation of the rotor, are interrelated in ac cordance with the following relations:

wherein the individual symbols have the meaning explained hereafter.

On the accompanying drawing:

FIGURES l-3 are explanatory diagrams illustrating the geometrical relations considered by this description;

FIGURE 4 is a cross-sectional View of a compressor according to this invention, and

FIGURE 5 is an axial sectional view on line VV of FIG. 4.

It should be preliminarily stated that all considerations set out in the appended description apply exclusively to compressors of the type specified and do not in the least apply to liquid pumps even if they are of a structure of the type specified.

Referring first to FIGURES '4 and 5, the compressor shown therein comprises a hollow stator 10, in the cylindrical cavity of which a cylinder 11 is sealingly rotatable, the geometrical axis of the cylinder being denoted by O. The stator 10 has end plates 12, 14 carrying antifriotion bearings 15, 16, respectively, centered on the axis 0. The cylinder 10 has two end discs 17, 18 bolted thereto, rotatably supported by their respective bearings 15, 16. The disc 18 is fast with a shaft 19 extending outwardly beyond the end plate 14 on the stator, by which the cylinder 11 can be rotated for operation of the compressor.

The end plate 12 on the stator serves as a support for a stud shaft 20, the axis 0 of which extends parallel with the axis 0 of the cylinder 11. The eccentricity of the two axes O, O is denoted by e. The stud shaft 20 is securely fixed to the plate 12 and extends from the latter through the end disc 17 on the cylinder towards the opposite disc 18 and ends by a journal 20' centered on the axis 0 of the cylinder and supported by an antifriction bearing 21 encased in the disc 18.

A cylindrical rotor 22 is freely rotatable on the stud shaft 20, extends coaxial with the axis 0' and is enclosed by the cylinder 11. In the construction shown the rotor 22 is formed with seven radial angularly equi-distant slots in which vanes P, P P (compare also FIG. 3) are slidably arranged. One of these vanes, denoted by P, is somewhat longer than the remaining ones and engages by its ends in diametrical grooves, formed in circular brass bosses 23, 24, respectively, rotatable in cylindrical cavities 25, 26 in the respective end discs 17, 18. The driving vane P therefore couples the rotor 22 with the cylinder 11 so that on actuation of the shaft 19 the cylinder and rotor concordantly rotate about their respective axes O, O. The diameter of the rotor 22 and its eccentricity e are so selected that the rotor is constantly tangential to the inner surface of the cylinder along a tangency generatrix 0 The axes O, O and tangency generatrix 0 are co-planar and are situated on a radial plane S which shall be taken hereafter as an angular reference plane. The vanes P P subdivide the crescent-shaped space between the rotor 22 and cylinder 11 in variable volume chambers C C Each of the chambers C C has associated therewith a slot A A respectively, formed in the cylinder 11 parallel with the generatrices of the latter. However, in the simplest case each of said slots can be reduced simply to a circular hole radially bored in the cylinder. Finally, the stator 10 has formed therein induction and delivery ports denoted by 28, 30, respectively. The intake port extends over an arc of about l40 and is at any rate so arranged as to afford best possible filling of the chambers -C C through their respective slots A A during operation. However, as far as the delivery port is concerned, the terminal edge 30b of the latter is close to the angular reference plane S. The leading edge of this port is denoted by 30a, its angular 0 position (advance angle") to the plane S representing one of the fundamental factors for the purposes of the invention. This advance angle is indicated by v in FIG. 1 and is taken at the center of the cylinder 11.

In operation, on starting from plane S the chambers C C successively increase in volume and draw gaseous fluid from the port 28 through their respective slots A A whereby they successively decrease in volume, so that the fluid trapped therein is compressed. The compression ratio A depends upon various factors, inter alia upon the advance angle of the delivery port, it being obvious that, as soon as one of the slots A A reaches the leading edge 363a of the delivery port 30, the fluid trapped and compressed in the respective chamber C C is transferred at constant volume from the chamber to the port 30 through the respective slot.

FIG. 1 considers a generic chamber C confined between vanes P and P and having associated therewith a slot A the said chamber being shown in its largest volume conditions. The fluid trapped therein is compressed till the slot reaches the position A' wherein its leading edge (looking in the direction of rotation F of the rotor and cylinder) coincides with the leading edge 30a of the delivery port 30. Overflow of the fluid to the delivery port 30 starts henceforth. It will therefore be clear that the performance accomplished by the generic chamber C is made up of two parts: substantially adiabatic compression (up to position P and A' of the blade P and slot A and constant volumes compression (overflow), the former only of which is related to the compression ratio which should be reached in each chamber, whereas the latter represents the energy expended for overflow. It should be moreover borne in mind that the specific adiabatic work (applied to an initial unit gas volume) is substantially less than the specific constant-volume compression work.

It should further be noted that in operation the position of a generic vane P with respect to the cylinder 11 is variable. Through the first 180 of rotation of the cylinder starting from the tangency generatrix 0 a vane P lags with respect to the cylinder and forms with its respective cylinder radius an angle which varies continuously in accordance with the law: zero-positive maximum-zero. On rotation through the subsequent 180 the blade P is accelerated with respect to the cylinder and forms with its respective radius an angle variable in accordance with the law: zero-negative maximumzero.

The invention starts from the concept that, when a generic slot A (FIG. 1) reaches the position A its respective vane P should reach a position P which, consistently with all further conditions to be set out hereafter, is as near as possible to the said slot, as well as from the concept that the above mentioned rule should apply to all vanes present including the driving vane. In otehr words, once a compression ratio A is established, each of the chambers C C should attain exactly this ratio at the moment its respective slot A A is set in communication With the delivery port 30. This implies that, adversely to known constructions, the angular distribution of the vanes P P and slots A A should fulfil a defined mathematical relation taking care of all factors involved. Among the latter, some can be considered as starting factors, such as, for instance:

anna Needless to say the thickness s of the vanes depends upon all the remaining dimensions and pressures involved, in that it defines the mechanical strength of the vanes. On the other hand, the circumferential width of the slots depends upon the volume delivery of the compressor; when circular apertures are employed for the slots the value a will be equal to the diameter of the aperture. As far as the driving vane P is concerned, its pivotal axis can be situated on its longitudinal middle line (this being the case shown on the drawings) or, for instance, on its outer edge or in any other position.

Starting from the data given above, the advance angle v of the delivery port 3% is first determined, the said angle being defined primarily on the basis of the reduction in volume to be undergone by a generic chamber C (FIG. 1) before being set in communication with the delivery port 30. Secondly, it will be seen that on further rotation in the direction F starting from the conditions shown in FIG. 1, the vane P' gets still nemer the slot A' and could even reach beyond the latter before having covered the angle 7, that is, before having accomplished its fluid-transfer work. When overflow begins each vane should therefore follow its respective slot at an angular distance 0' Which, according to this invention, amounts to:

wherein A" amounts as an average to about 5 and never exceeds 10. The accurate value of A can be reckoned or found experimentally for each vane, but calculation is so tedious and dilferences so negligible in practice that a A=5 is conveniently adopted with all vanes.

The total angle:

defines the position of the contact generatrix between any vane and the cylinder 11 when its respective slot is about to communicate with the delivery port 30; On assuming A=5 (constant for all vanes), the angle f is likewise constant for all vanes and should be considered positive The situation shown in FIG. 2 is now assumed, where in the slot A (associated with vane P is about to be placed in communication with the delivery port. The vane P forms with the driving vane P an angle a taken at the center 0' of the rotor 22 and contacts the cylinder 11 along the generatrix N of the latter which is spaced by the angle from the tangency generatrix O Indicating by M the pivotal axis of the driving vane, the leading edge of the slot A forms at this moment with the radius OM=r an angle {3 Moreover, at the same time, the vane P forms an angle 7 with the radius ON=R, the driving vane forming an angle A with the radius OM :1; Since, as mentioned above, each of the vanes P P with its respective slot should reach in turn the condi tions of vanes P shown in FIG. 2, the angle 7 is constant for all vanes P P The value of this angle is obtained from the triangle OO"N, from which:

7=arc to e. am i wherein the three bracketed angles are all positive. For the vane P The angle 6 therefore amounts to:

wherein 'y and u always are positive while 6,, is considered with its sign resulting from the general formula:

6 =are sin sin (fin-bard) which corresponds to the particular Formula '6.

Therefore the Formula 10 serves for calculating for instance the angles ,6 of the slots A A when the vane distribution angles oa are given, and vice-versa. The angles oa and ,3, are azimuth angles taken at the centers and 0, respectively, in the direction of rotation of the rotor starting from the tangency generatrix 0 when the driving vane P is situated on the plane OOO In the diagram shown in FIG. 3 the vanes P P are assumed to be angularly equidistant, so that the angles a a etc., take the values 5125, 10250 etc., respectively. By substituting these values in the Formulae l0 and 11 the corresponding angles 5 B etc. for their respective slots A A A are obtained. The compressor shown in cross sectional view in FIG. 4 has a distribution of its slots and vanes resulting from Formulae l0 and 11 by establishing:

wherein k is a whole number between 1 and (nl). If

desired, a reverse situation can be considered, wherein the slots are angularly equidistant by establishing 360 aw n 13) and wherein the azimuth angles of the vanes P P will be calculated according to Formulae and 11.

The position of the driving vane does not require any calculation, in that it has been taken as the origin in the above considerations. The position of its respective slot A however is still to be determined. This is readily accomplished by substituting in FIG. 2 the driving vane P for the vane P it is then found that the azimuth angle of the slot A should necessarily equal:

B=a= -57.29+A 14 in accordance with Formula 1.

The mutual distribution of the slots and vanes according to this invention significantly differs from a uniform distribution heretofore usually adopted in compressors of the type specified, as shown by the following table:

Table 1 Azimuth angle Slot slots dis- Diflerence equidistant tributed slots according to the invention 12047 l1326 721. 17704 16453'... 1258. 23l51. 2l620 31. 27953. 26747. 1206. 32443- 31914- 529. 1035- 1032'- 003.

However, practical difierences between a compressor of a construction according to this invention and a compressor having uniformly distributed slots and vanes R'=75.3 mm. s=4.5 mm. r=60.3 mm. =5'8 e=8.3 mm. n=7 a=3.5 mm.

The axial length of the chambers C C was assumed to be 15 cm.; suction pressure=1 kg./sq. cm. abs. pressure; specific volume of air drawn at 15 C. =0.82 mfi/kg.

Table 2 Equidistant arrange- Arrangement according ment of slots to the invention Chamber Pmax evee, mux eve gauge, Mi, kgmJ gauge, Mr, kgm./ kgJsq kgm. kg. kg./sq. kgm. kg.

cm. cm.

In this Table P denotes the gauge pressure in the individual chambers on commencement of overflow, M the corresponding opposing torque, and LSpec the specific work accomplished. Table 2 clearly shows non-uniformity of operating conditions in the various chambers in known compressors as compared with full uniformity afforded by the compressor according to the invention. In the known compressor some chambers do not even reach the operational pressure (5.8 kg./sq. cm.) and prevailingly perform compression work at constant volume thereby absorbing enormous power quantities. Further chambers, such as C and C accomplish a quite useless over-compression work. The opposing torque varies periodically between 5.75 and 9 kgm., while the specific work likewise periodically varies between 13,300 and 37,170 kgm./kg. Finally, by summing up the specific works it will be seen that, apart from all frictional works, the known compressor absorbs almost twice the power absorbed by the compressor according to this invention.

In calculating the specific works tabulated in Table 2 the following formulae were adopted:

1' (P2P1) which apply to the adiabatic compression and constant volume compression, respectively, wherein m is the exponent of the polytropic curve.

What I claim is:

A slidable vane rotary compressor comprising; a stator having a cavity, a cylinder rotatable in the cavity of the stator, a cylindrical rotor rotatable within the cylinder about an axis parallel to but offset from the cylinder axis in an arrangement such that the rotor is constantly tangential to the cylinder along a contact generatrix, a number of vanes accommodated by and slidable in the rotor in a substantially radial direction including a single driving vane pivotally connected to the cylinder about an axis parallel with the axes of the cylinder and rotor thereby mutually coupling the rotor and the cylinder for rotation in the stator, a fluid-transfer aperture in the cylinder in a location between each pair of vanes, and induction and delivery ports in the stator with which said apertures connect in succession on rotation of the rotor and cylinder; said compressor being characterized by; a mutual arrangement of the vanes, fluid-transfer apertures and delivery port wherein, considering the driving vane on the contact generatrix and taking as angular reference a radial plane extending through said generatrix and axis of the cylinder, the azimuth angles oa of the remaining vanes about the rotor axis and the azimuth angles B of their respective fluid-transfer apertures about the cylinder axis, measured in the direction of rotation of the rotor, are interrelated according to the relations:

wherein a=circumferential width of fluid-transfer apertures s=vane thickness e=eccentricity of the rotor R=inner radius of the cylinder r=pivotal radius of the driving vane n=nunrber of vanes K=an integer ranging between 1 and (n1) zadvance angle of the delivery port, and A =angle not exceeding 10.

References Cited in the file of this patent UNITED STATES PATENTS 1,253,716 Palmer Jan. 15, 1918 1,610,748 Cozette Dec. 14, 1926 1,787,543 Nichols Ian. 6, 1931 2,018,391 Whitfield Oct. 22, 1935 FOREIGN PATENTS 138,395 Switzerland May 1, 1930 582,395 Great Britain Nov. 13, 1946 929,387 Germany June 27, 1955 

